Variable ratio gearmotor with interactive ratio control

ABSTRACT

A variable ratio gearbox is combined with a drive motor and an interactive control system. A coaxial shaft connects to a differential gear set to power an output shaft. The transmission has a drive shaft with an inner shaft and an outer shaft. A friction disk nonrotatably mounts to said outer shaft and a shifter connected to the friction disk moves the friction disk along the outer shaft. At least two cones engage the periphery of the friction disk and a differential gear set combines the rotation of the inner and outer shaft.

BACKGROUND OF THE INVENTION

The disclosure pertains to geared drives commonly referred to as “gearmotors”. Gear motors drive a wide range of different industrialmachines, such as pumps, conveyors, rock crushers, rotary kilns, hoists,and some types of vehicles.

Many of these types of machines, such as oil well pumping units, airconditioning compressors, loaded ore conveyors, sand and gravelconveyors, loaded rotary kilns and loaded hoists require relatively highdriving torque for starting and bringing up to operating speed.Typically, such machines are started and driven by electric motorsacting through fixed ratio gearing and are therefore subjected to highstarting loads. Hence, to handle the starting loads, such driving motorsare usually sized larger than is required for steady running. It is wellknown that electric motors are most efficient when sized to be near fullload during steady operation. Thus, the high starting loads combinedwith less than optimum running efficiency causes considerable waste ofelectric power.

Also, some of these types of machines, such as oil well pumping units,rock crushers, hoists and the like, experience widely varying torqueloads during normal operation. Typically, electric motors used fordriving such varying loads are the NEMA D “high-slip” type in order towithstand the load variation without overheating or having to beextremely oversized. Of course, both “high-slip” motors and oversizedmotors are considerably less efficient than correctly sized “low-slip”or “premium efficiency” motors.

In addition, some of these types of machines, such as oil well pumpingunits, gas compressors, boiler feed pumps and hoists must operate atvariable output speeds to accommodate changing operating requirements.Typically, such speed changes are accomplished by electronic speedvariation of the drive motor or by hydraulically varying the speedrelation between the drive motor and the load. Both of these methodsproduce undesirable inefficiencies.

Therefore, it is an object of the invention to provide a gear motorhaving a suitably variable mechanical ratio for driving a wide varietyof different machines at suitable speeds and/or variations of speeds ina highly efficient manner; to reduce and/or eliminate inherent motorstarting loads; and to facilitate the use of premium efficiency motorsinstead of other, less efficient types.

SUMMARY OF THE INVENTION

The invention is a combination of an infinitely variable ratio gearboxwith an drive motor and an interactive control system. A coaxial shaftconnects to a differential gear set to power an output shaft.

The gearbox is arranged to provide infinite ratio at zero output speedwhereby the drive motor can be started without load. The control systemadjusts the gearbox ratio in a continuous or “step-less” manner inresponse to a motor load ampere signal so as to provide a smooth,controlled start up and controlled operation of a load. The controlsystem maintains the drive motor at its most efficient speed and loadwhile continuously adjusting the gearbox ratio for optimum output torqueand speed for the particular load application.

Output speed can be adjusted either by direct shift of the gearbox ratioor by adjusting the motor ampere load reference signal. For instance, ahigher ampere setting will produce higher output acceleration and/orspeed while a lower ampere setting will produce lower outputacceleration and/or speed. The control system may be arranged forautomatic adjustment of the ampere reference signal and/or automaticstart up or shut down of the system.

When used in electric powered vehicles, the invention can enable higheroutput torque at low speeds and higher maximum output speeds for thegiven drive motor size than conventional methods.

For industrial applications, the drive motor is preferably a poly-phasealternating current motor of the highest efficiency design and ismaintained at its most efficient speed. Other types of electric motorsmay be used for particular applications. For example, direct currentmotors may be preferred for use in vehicles, especially for batterypowered vehicles.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a partially sectioned side view of an apparatus formedaccording to the present invention;

FIG. 2 is a hydraulic schematic illustrating the hydraulic circuit of anembodiment of the invention;

FIG. 3 is an electrical schematic illustrating the power circuits andcontrol circuits of an embodiment of the invention;

FIG. 4 is a schematic illustration of the control module for anembodiment of the invention;

FIG. 5 is a partially sectioned side view of an alternate embodiment ofthe invention; and

FIG. 6 is a cross sectional view of an alternative embodiment of thecone.

DETAILED DESCRIPTION OF THE INVENTION

In FIG. 1, an embodiment of the invention is shown in a typicalenvironment wherein the invention is used to drive a machine such as anoil well pumping unit.

The transmission has a central tractionally driven member or disk 1mounted upon and with a hollow supporting coaxial shaft 2. All elementsof the apparatus are located within a housing 3 and cap 15. The shaft 2is supported on appropriate bearings, such as bearing set 4 at the frontend of shaft 2 and an appropriate bearing at its opposite end mountedwithin the hub of a planet gear carrier 28. Disk 1 is nonrotatablymounted upon a concentric hub 5 by any conventional means such as aspline conforming to and engaging with a straight spline portion ofshaft 2. Hub 5 is so arranged to allow disk 1 to travel axially alongshaft 2 in a low friction manner while simultaneously transmittingtorque cooperatively. Although disk 1 is shown as a single disk, it maybe of a multiple disk structure.

A shifting collar 6 connects to hub 5 by means of a bearing such as athrust bearing within collar 6. Collar is arranged to move axially alongshaft 2 to thereby control axial position of disk 1 while allowing freerotation of disk 1 and shaft 2 relative thereto.

An actuator controls the axial position of hub 5 and disk 1. FIG. 1depicts one type of actuator having a floating lever 116 pivotallymounted to collar 6 and pivotally anchored by a fulcrum link 117 whichis in turn pivotally fastened to cap 15 generally as shown. The drivenend of lever 116 is pivotally attached to actuating rod 51 of a servoactuator 50. As shown herein, actuator 50 is electrically driven, butcould be hydraulically driven.

A plurality of conical rotors 7 are symmetrically positionedcircumferentially about disk 1 so that the inwardly facing sides of thecones 7 are parallel to shaft 2 and in frictional engagement with therim of disk 1. In the present embodiment, five cones 7 may be providedabout the periphery of disk 1, though only two are shown. Fortractionally driven elements such as disk 1, the fatigue life of thedriven element as well as the bearings on which they may be supported,can be calculated using the following commonly known formula:

Life (hours)=K/RPM×(Rated Load/Applied Load)³

From this, it should be noted that the applied radial, or normal,traction contact load, which is proportional to torque load, must beheld to relatively light limits to prevent rapid fatigue failure.Further, applied load has a much greater effect on fatigue life thandoes rotational speed of the elements. The torque output from disk 1 isalso proportional to the applied traction contact normal load and numberof tractional contact points. Again, the optimum number of tractionaldrive members or cones 7 may be therefore chosen for the particularenvironment and application of the power transfer system. In the presentinvention, the preferred embodiment utilizes the maximum number ofcontact points to provide maximum power output, with final output speedpredetermined by a differential gear assembly which will be describedhereafter. The disk member 1 as well as the tractional drive member orcone 7 are designed and arranged to sustain extremely high rotatingspeeds without causing premature failure. The cones 7 of this embodimentare fifteen degree (15°) cones each having approximately four to one(4:1) diametrical ratio, although other embodiments might include conesof a different angle, size and ratio configuration. In this embodiment,the major outside diameter of each cone 7 is approximately equal to theoutside diameter of disk 1. U.S. Pat. No. 6,001,042, incorporated hereinby reference, shows the cone assembly in more detail.

Each cone 7 includes a concentric shaft extending from each endsupported by bearings 8 and 9. A drive gear 10 is provided at its frontend, with a thrust bearing 12 supporting the shaft at its front end, asshown. Bearings 8 and 9 are, preferably, needle-type roller bearingscapable of sustaining extremely high rotating speeds and relatively highradial loads. Bearings 8 and 9 are mounted in bearing blocks 13 and 14,respectively, which are in turn fastened to a cone assembly cap 15attachable to the main housing 3. A concentric piston 16 is arranged toabut thrust bearing 12 rearward and to be sealed by means of an o-ring17 within a forcing cylinder 18. The cylinder 18 in turn abuts the endplate 23 of cap 15 forwardly so that piston 16 applies axial thrustthrough bearing 12 to cone 7 whenever fluid pressure is directed througha port 19 into cylinder 18 and against piston 16. In this manner,clamping force, or normal force, between cones 7 and disk 1 is achievedto prevent slippage. Bearings 8 and 9, and the mating shaft journals ofcone 7, are arranged to allow slight but adequate axial movement of cone7. Drive gear 10 may be keyed on its shaft and is retained on its shaftby axial force through bearing 12.

As shown in FIG. 1, an input ring gear 21 is provided in common meshwith drive gears 10 at a step-up ratio of 1:5 and is supported inbearing set 22 in association with end plate 23. Bearing set 22 ispreferably of the type capable of supporting both radial and axial loadsand is retained within plate 23 by a retaining plate 24 and by nut 26which is in turn installed on the hub of gear 21. Bearing set 4comprises a pair of opposed thrust bearings arranged to support andcontrol axial force and movement of shaft 2 and is mountedconcentrically within gear 21. Gear 21 is arranged to be coupled to anddriven by a suitable driver, such as electric motor 53.

Drive motor 53 may be a flange mounted type which may be mounted toplate 23 and cap 15 by means of an adapter ring 52. A sleeve typecoupling 25 is keyed to the shaft 54 of drive motor 53 and is splined tothe hub of gear 21 and to the end of an internal coaxial shaft 27. Theshaft 27 is provided in common mesh with coupling 25 and extends throughcoaxial bores in other elements of the mechanism rearward where itmeshes by splines with planet carrier 28.

A set of differential gears combines the torque of the shafts 2 & 27.One example of the gear set comprises output planet gears 43 which spinon bearings mounted on axially disposed spindles secured to carrier 28.Planet gears 43 are in common mesh with sun gear 45 which is in turnkeyed onto shaft 2. Ring gear 44 is also in common mesh with planetgears 43 and the hub of gear 44 rotates on a bearing mounted on the neckof carrier 28. The hub of gear 44 is supported in bearing set 46 and isintegral to output shaft 47. Shaft 47 will of course be coupled to drivea load. In this embodiment, the diametrical ratio between ring gear 44and sun gear 45 is 4.5:1, although other ratios are of course possible.

Parallel or “split torque” transmission of torque is accomplished by thetransmission of one torque reaction through gear 21, gears 10, cones 7,disk 1, shaft 2 and sun gear 45 to the differential planet gears 43; andthen another torque reaction through shaft 27, through carrier 28 toplanet gears 43, which is the other parallel torque path. Thedifferential action of planet gears 43 combines the two separate torquesand directs a single torque to ring gear 44. In this manner, the maintorque load is transmitted through shaft 27, while the torque loadthrough cones 7 and disk 1 is minimized, yet facilitating the variationof ratio by axial movement of disk 1. A position transducer 88 isattached to the pivot end of rod 51 to provide a ratio position signalfor disk 1.

Although not shown herein, the proposed load, being an oil well pumpingunit, may include a spring applied and electrically releasable brake foraiding in safely stopping operation whenever power is disconnected.

Also, as shown in FIG. 5, any conventional clutch 90 may be installedwithin gear 44 and between gear 44 and carrier 28 to selectively connectgear 44 to carrier 28 in certain applications of the invention. Clutch90 may be a commercially available type with known characteristics andengagement methods, so its details are not shown here. Its usage will befurther explained later.

FIG. 2 schematically illustrates a hydraulic system for supplyingcontrolled pressure to pistons 16. The pressure system comprises a pump55 which would, preferably, be driven by a small drive motor separatefrom main motor 53. Pump 55 would, preferably, also be a variabledisplacement type to minimize energy consumption. Pump 55 will intakehydraulic fluid from a sump 60 and supply pressure through line 56 to apressure control valve 89. Gauge 57 is also connected to line 56 to readout pressure. Valve 89 provides controlled pressure through line 87through shut off valve 86 to line 58 which is connected to all ports 19and thus to pistons 16. Shut-off valve 86 should be the normally closedtype which opens when energized and closes when de-energized. Also,valve 86 should be equipped with an adjustable bleed-off means so thatit can bleed-off pressure in a desired manner. A pressure switch 81 isinstalled in line 58 to indicate when pressure is up for operation andgauge 59 is connected to line 58 to read out pressure. Valve 89 has ausual drain line into sump 60.

In addition, although not shown herein, a lubrication system will berequired to both lubricate and cool various moving components, such asbearings, gears, cones and disks. Since such a lubrication system ormethods are commonly known by those skilled in the art, it will not beshown herein.

FIG. 3 schematically illustrates both the power circuits and controlcircuits relating to the systems previously described. Transformer 77converts the high power line voltage to 120 volt AC control power. Mainmotor and pump motor 55 are shown connected to the power circuit. Also,a bridge rectifier 84 is connected to the power circuit to supply DCpower to a brake coil 85. Usage of brake coil 85 will be explainedlater.

Pump motor 55 is controlled by a 3-phase starter/contactor whichcomprise the usual coil 126, power contacts 126 p and control contacts126 a 1, 126 a 2 and 126 a 3 as shown.

Main motor 53 is controlled by a Y/Delta starting system which comprisesa main contactor having the usual coil 123 and associated power contacts123 p and control contacts 123 a 1, 123 a 2 and 123 a 3; a Y contactorhaving the usual coil 120, associated power contacts 120 p and anormally closed control contact 120 a; and a DELTA contactor having theusual coil 121, associated power contacts 121 p and control contacts 121a and 121 c. The pump starter/contactor 126 and the main contactor 123include safety overload tripping contacts 78 and 79, respectively.

A neutral position switch 80 is located in the control circuit so thatcontactor 123 cannot energize unless actuator 50 and disk 1 are atneutral position. Based on the various component ratios presentedherein, the neutral position for disk 1 is near the large end of cones 7which yields a combined ratio of the mechanism equal to infinity andthus an output speed at shaft 47 equal to zero. Conversely, for lowestcombined ratio and maximum output speed at shaft 47, the position ofdisk 1 will be near the small end of cones 7.

A timing relay 124, having associated delay contacts 124 od and 124 cd,provides a delayed transfer from Y motor connection for starting to aDELTA motor connection for running. For instance, when main contactor123 is energized, timer 124 will be energized through the auxiliarycontact 123 a 2 as shown. Also, at this time Y contactor 120 will beenergized through the normally closed contacts 124 od and 121 a. Thepower contacts 120 p will connect motor 53 for Y configuration for lowampere starting. Timer 124 will be set to time out after motor 53 hasachieved full speed so that Y contactor 120 will be de-energized and Dcontactor will 121 be energized through auxiliary contacts 124 cd and120 a, as shown. The power contacts 121 p connect motor 53 for DELTAconfiguration for full power running. Contactors 120 and 121 aremechanically interlocked, as denoted by hidden lines, so they cannotboth be energize at the same time.

The starter coil 123 is accompanied by a timing relay 125 equipped withboth a delay contact 125 d and an instant contact 125 i. In a startingsequence, relay 125 will be energized first and its instant contact 125i will energize brake coil 85 before contactor 123 energizes motor 53.After coil 85 has sufficient time to begin releasing the brake(approximately 0.5 second), the delay contact 125 d will energizecontactor 123 which then energizes motor 53 and the subsequent Y/DELTAsequence. Actually, Y/DELTA starting systems are well known to thoseskilled in the art.

The control circuit may also be equipped with a pilot relay 127 havingan auxiliary contact 127 a and arranged to be energized through a masterON-OFF switch 128. The circuit may also include a MANUAL START switch129 and a MANUAL STOP switch 130.

An auto-start contact 75 and an auto-stop contact 76 may be included forautomatic starting and stopping the system. Contacts 75 and 76 will belocated remotely in a conventional “pump-off control unit” normallyfound on oil well pumping units. Also, the pressure switch 81 previouslymentioned is included in the circuit to prevent start up of main motor53 until oil pressure and lubrication is up to a preset amount. Inaddition, thermostatic switches 82 and 83 are included to preventrunning of main motor 53 when the oil temperature is not within presetlimits.

During any start up sequence, the gearbox ratio is infinity and thetorque at shaft 47 is zero until motor 53 reaches operating speed and isswitched from Y connection to DELTA connection. Thus, motor 53 alwaysstarts at zero load.

A set of current pick up coils 92 are installed on the power conductorsto motor 53 to provide a motor load signal through line 68 to thecontrol system to be explained later herein.

FIG. 4 schematically illustrates a control module and its associatedcircuits. Module 70 includes a programmable microprocessor havingmultiple input and output signal capability. Module 70 also includes aservo drive power supply controllable by the microprocessor.

The face of module 70 may include an ammeter 71 to provide read out ofmotor load based on the load signal provided through line 68. Module 70is also equipped with a selector switch 72 for selecting between on/offoperating mode and slow-down/speed-up mode. For instance, in on/offmode, a pumping unit will be stopped when a pump-off condition issignaled and restarted automatically by a time clock or the like. Inslow-down/speed-up mode, the pumping unit will merely slow down a presetamount when a near pump-off condition is detected and speeded up apreset amount after a preset amount of time to keep the pumping ratejust below well production rate.

Module 70 is also equipped with a manual adjustment 73 for setting theampere rate which controls the servo driver. In addition, module 70 isequipped with a manual adjustment 74 for setting the rate of ampere riseduring startup, which determines the output starting torque andacceleration. Adjustment 74 will be set to provide an optimumacceleration rate which does not overload the mechanism of the gear boxor the mechanism of the pumping unit.

Module 70 receives seven input signals and produces one output signalplus output servo power. For instance, voltage through normally closedcontact 121 c through line 69 signals module 70 to position servoactuator 50 for zero motor load and to an infinite ratio or “neutral”position. Servo drive power is directed through line 62 while theservo/ratio position signal is received through line 64.

Of course, when motor 53 is running, the normally closed contact 121 cshown in FIG. 4 is open so that the servo driver can position actuator50 according to the ampere load setting. Thus, based on the motor loadsignal received through line 68, the servo driver will continuallyposition actuator 50, and thus the ratio of disk 1, so as to maintainthe set motor load. During this time, based on the combined signals fromline 64 for ratio position, from line for oil temperature and from line68 for motor load, clamp pressure signal through line 63 is modulated tomaintain optimum pressure to pistons 16 to prevent slippage between disk1 and cones 7 at the prevailing torque load.

Such applications such as an oil well pumping unit experience drasticload fluctuations during each cycle of the pumping unit duringoperation. Hence, in working to keep motor load at the set ampere rate,the servo driver of module 70 will continually position actuator 50 anddisk 1 in phase with the changing torque load of the pumping unit. Forinstance, as torque load at shaft 47 increases, disk 1 will becontinually shifted to a higher ratio as shaft 47 slows and the torqueload on motor 53 holds steady. As torque load at shaft 47 decreases,disk 1 will be continually shifted to a lower ratio as shaft 47 speedsup and the load on motor 53 remains steady. Although the pumping unitwill speed up and slow down during each stroke cycle, a given motorampere setting will produce a given overall stroke rate. The motorampere setting is increased to increase the stroke rate and the amperesetting is decreased to reduce the stroke rate.

When operating in the slow-down/speed-up mode, selector 72 will be inthe slow-down position. The pumping unit will normally be equipped witha conventional “pump-off controller” (not shown) which can supply anauto ampere decrease signal through line 66 whenever a near pump-offcondition is detected. Module 70 will be pre-programmed to automaticallyreduce the ampere setting a preset amount whenever an auto amperedecrease signal is received. Thus, the pump stroke rate will be reducedso as to not pump-off the well, which might cause pump cavitation andpossible damage. In addition, the conventional “pump-off controller” maybe equipped with a timing device to periodically send an auto ampereincrease signal to module 70 which may be programmed to slightlyincrease the ampere rate and thereby the stroke rate. In this manner,the stroke rate may be adjusted to an optimum rate without causing anactual pump-off condition.

When operating in the on/off mode, selector 72 will be in the on/offposition. In this mode, whenever the conventional “pump-off controller”detects a near pump-off condition it can send an auto stop signalthrough line 91 to module 70 which then initiates a pump shut downsequence. In a shut down, the auto stop contact 76, shown in FIG. 2,(included in the “pump-off controller”) will open which de-energizes.contactor 126 which in turn de-energizes contactors 123 and 121 andrelay 125. Contact 126 a 3 opens to de-energize valve 86 when contactor126 is de-energized. Thus, valve 86 closes to hold clamping pressure onpistons to prevent slippage between cones 7 and disk 1 during shut down.Valve 86 should be adjusted to allow sufficient bleed off of pressureafter shut down so that disk 1 can safely move axially without rotating.Also at this time, 125 i opens to de-energize brake coil 85 setting thebrake on the pumping unit. The brake will be adjusted to bring thepumping to a soft stop and prevent roll-back. Also, during this time,contact 121 c shown in FIG. 4 closes to provide a signal through line 69which directs module 70 and the servo driver therein to positionactuator 50 and disk 1 to neutral position. Thus, the pumping unit,motor 53 and the entire mechanism will be brought to a stop.

Likewise, in the event of an unplanned shut down, such as loss ofelectric power during a lightning storm, shut down will occur in thesame manner as previously described, except that module 70 and the servodriver will not return disk 1 to neutral position until electric poweris restored.

To automatically restart, the conventional “pump-off controller” willinclude the auto-start contact 75, shown in FIG. 2, which may close at apreset time. Start up is thus achieved in the manner previouslydescribed.

In some applications, other than and unlike oil well pumping units, theload needs to run at a steady speed when up to speed and in operation.In such steady speed applications, an arrangement such as clutch 90shown in FIG. 5 previously mentioned, can be utilized to connect carrier28 to the hub of gear 44 after gear 44 has been brought up to operatingspeed. In this manner, the driving torque from motor 53 can be directlytransmitted through shaft 27, to carrier 28, through clutch 90 to ringgear 44, and out through output shaft 47 without need of any torque loadon disk 1 during steady running.

In order to utilize clutch 90, the ratios of gears 21, 10, 45, 43 and44, as well as the cone-to-disk ratio, must be arranged so that whendisk 1 reaches the small end of cones 7 during acceleration, the speedof ring gear 44 is equal to the speed of carrier 28, which enablessmooth engagement of clutch 90.

When clutch 90 is thus engaged, module 70 and the servo driver may beprogrammed to keep disk 1 at a synchronized position which preventstorque application to disk 1. Also, during this time the clampingpressure through line 58 to pistons 16 can be reduced to a minimumamount whereby disk 1 and cones 7 and associated bearings are notsubjected to any consequential stress or wear during steady operation ofa load.

An alternative cone, depicted in FIG. 6, has a circumfrentiallyextending lightening groove 110 formed in the end face of the cone andcommunicating with lightening holes 111. While any number of holes maybe used, it is envisioned that each cone will have six. The lighteningholes 111 meet at a common point and a cooling oil passage 112 extendsfrom this point to the opposite end face of the cone.

While the invention has been described with reference to preferredembodiments, a wide range of sizes, variations, alterations ormodifications of the invention are possible without departing from theintent and scope of the invention.

1-19. (canceled)
 20. A transmission comprising a drive shaft, said driveshaft having an inner shaft and an outer shaft, a friction disknonrotatably mounted to said outer shaft, a shifter connected to saidfriction disk, said shifter comprising an actuating system to causemovement of the friction disk along the outer shaft, at least two conesengaging the periphery of said friction disk, and a differential gearset combining the rotation of said inner and outer shaft.
 21. Thetransmission of claim 1, further comprising an output shaft connected tosaid differential gear set.
 22. The transmission of claim 1, including aclutch for connecting said inner shaft to said output shaft.
 23. Aninfinitely variable ratio gearbox comprising a drive motor and a controlsystem for operation of a transmission having a coaxial shaft having aninner and outer rotatable shaft connected to a differential gear set tocombine the rotation of the inner and outer shaft to power an outputshaft connected to a load, wherein the transmission is arranged toprovide infinite ratio at zero output speed whereby the drive motor canbe started without load.
 24. The infinitely variable ratio gearbox ofclaim 4, wherein the control system is configured to adjust the gearboxratio in a continuous manner in response to a motor load ampere signalsupplied to the control system so as to provide a smooth, controlledstart up and controlled operation of the load.
 25. The infinitelyvariable ratio gearbox of claim 4, wherein the control system maintainsthe drive motor at substantially its most efficient speed and load whilecontinuously adjusting the gearbox ratio for optimizing output torqueand speed for the load for a given application.
 26. The infinitelyvariable ratio gearbox of claim 5, wherein the control system controlsthe output speed of the drive motor either by direct shift of thegearbox ratio or by adjusting the motor ampere load reference signal.27. The infinitely variable ratio gearbox of claim 5, wherein thecontrol system controls the output speed of the drive motor by adjustingthe motor ampere load reference signal, wherein a higher ampere settingwill produce higher output acceleration and/or speed while a lowerampere setting will produce lower output acceleration and/or speed. 28.The infinitely variable ratio gearbox of claim 5, wherein the controlsystem is configured to cause automatic adjustment of the amperereference signal.
 29. The infinitely variable ratio gearbox of claim 4,wherein the control system is configured to cause automatic start up orshut down of the system.
 30. The infinitely variable ratio gearbox ofclaim 4, wherein the drive motor is a poly-phase alternating currentmotor.
 31. The infinitely variable ratio gearbox of claim 4, wherein thetransmission comprises a drive shaft, said drive shaft having an innershaft and an outer shaft, a friction disk nonrotatably mounted to saidouter shaft, a shifter connected to said friction disk, the shiftercomprising an actuating system to cause movement of the friction diskalong the outer shaft, and at least two cones engaging the periphery ofthe friction disk, wherein the control system is configured tocontinually position the friction disk via the actuating system tosubstantially maintain a set motor load.
 32. The infinitely variableratio gearbox of claim 4, wherein the load is a pumping unit.
 33. Amethod of controlling drive to a load comprising providing an infinitelyvariable ratio gearbox comprising a drive motor and a transmissionhaving a coaxial shaft having an inner and outer rotatable shaftconnected to a differential gear set to combine the rotation of theinner and outer shaft to power an output shaft connected to a load,controlling the transmission to provide infinite ratio at zero outputspeed whereby the drive motor can be started without load.
 34. Themethod of claim 14, further comprising controlling the gearbox ratio ina continuous manner in response to a motor load ampere conditions so asto provide a smooth, controlled start up and controlled operation of theload.
 35. The method of claim 14, further comprising controlling thedrive motor to operate at substantially its most efficient speed andload while continuously adjusting the gearbox ratio for optimizingoutput torque and speed for the load for a given application.
 36. Themethod of claim 14, further comprising controlling the output speed ofthe drive motor either by direct shift of the gearbox ratio or byadjusting a motor ampere load reference signal.
 37. The method of claim14, further comprising controlling the output speed of the drive motorby adjusting the motor ampere load reference signal, wherein a higherampere setting will produce higher output acceleration and/or speedwhile a lower ampere setting will produce lower output accelerationand/or speed.
 38. The method of claim 14, further comprising controllingthe operation of the gearbox to cause automatic start up or shut down ofthe gearbox.
 39. The method of claim 14, further comprising providingthe transmission with a drive shaft having an inner shaft and an outershaft, a friction disk nonrotatably mounted to the outer shaft, ashifter connected to said friction disk, the shifter comprising anactuating system to cause movement of the friction disk along the outershaft, and at least two cones engaging the periphery of the frictiondisk, and controlling the position the friction disk via the actuatingsystem to substantially maintain a set motor load.